Pulley support double row ball bearing

ABSTRACT

A pulley support double row ball bearing is disclosed, which improves the seal performance of the seal rings and is able to maintain sufficient durability even under severe conditions. The ball bearing includes inner sections of the seal rings which overlap the end surfaces in an axial direction of the inner ring such that the width of the overlap in the axial direction is 30% or more than the diameter of the balls. Moreover, tip edges of first and second protrusions are formed on the inside surface of the seal lips and come in sliding contact with end surfaces of the inner ring. Furthermore, third protrusions come close to and face the corner sections of the inner ring to form labyrinth seals in those areas.

TECHNICAL FIELD

The pulley support double row ball bearing according to the presentinvention, for example, is built into automotive auxiliary equipmentsuch as a compressor constituting an automotive interior airconditioning apparatus, and is used for rotatably supporting a pulleyfor rotationally driving the automotive auxiliary equipment with respectto a fixed support member such as a housing.

BACKGROUND ART

For example, as a compressor for compressing refrigerant, which is builtinto a vapor compression type refrigerator built into an automotive airconditioning apparatus, conventionally several types of mechanism areknown. For example, Japanese Unexamined Patent Publication No. H11-280644 discloses a swash-plate type compressor which convertsrotational motion of a rotation shaft into reciprocating motion of apiston using a swash-plate, and performs compression of refrigerant bythis piston. FIG. 6 and FIG. 7 illustrate one example of such aconventionally known swash-plate type compressor.

A casing 2, constituting a compressor 1, is formed by sandwiching acentral main body 3 between a head case 4 and a swash-plate case 5 fromboth sides in the axial direction (left-right direction in FIG. 6), andthen joining these with a plurality of fastening bolts (not shown). Onthe inside of the head case 4, there is provided a low pressure chamber6 and a high pressure chamber 7. Also, between the main body 3 and thehead case 4, a tabular partition plate 8 is sandwiched. The low pressurechamber 6, which is shown in FIG. 6 as if divided into a plurality ofsections, has the sections communicating with each other and connectedto a single inlet port 9 (FIG. 7) provided on the outside surface of thehead case 4. Furthermore, the high pressure chamber 7 is connected to anoutlet port (not shown) also provided on the head case 4. Moreover, theinlet port 9 is connected to the outlet of an evaporator (not shown)constituting this vapor compression type refrigerator, and the outletport is connected to the inlet of a condenser (not shown) constitutingthis vapor compression type refrigerator.

Within the casing 2, a rotation shaft 10 in a state of spanning betweenthe main body 3 and the swash-plate case 5, is freely supported forrotation alone. That is to say, both ends of the rotation shaft 10 aresupported by a pair of radial needle bearings 11 a and 11 b, on the mainbody 3 and the swash-plate case 5, and the thrust load exerted on thisrotation shaft 10 is freely supported by a pair of thrust needlebearings 12 a and 12 b. Of the pair of thrust needle bearings 12 a and12 b, one (right hand side in FIG. 6) thrust needle bearing 12 a isprovided between a part of the main body 3 and a step portion 13 formedon one end (right end in FIG. 6) of the rotation shaft 10, via a discspring 14. Also, the other thrust needle bearing 12 b is providedbetween a thrust plate 15 externally fitted to the outer circumferentialsurface of an intermediate part of the rotation shaft 10 and theswash-plate case 5.

Moreover, on the inside of the main body 3 constituting the casing 2surrounding the rotation shaft 10, is formed a plurality (for example inthe example shown on the figure, there are six evenly spaced in thecircumferential direction) of cylindrical bores 16. Inside the pluralityof cylindrical bores 16 formed in such a way on the main body 3, asliding portion 18 provided at the tip half portion (right half of FIG.6) of the respective pistons 17 is fitted to allow free displacement inthe axial direction. Moreover, the space between the bottom face of thecylindrical bore 16 and the tip end surface of the piston 17 (right endsurface in FIG. 6) serves as a compression chamber 19.

Furthermore, the space which exists on the inside of the swash-platecase 5 serves as a swash-plate chamber 20. On the outer circumferentialsurface of the intermediate part of the rotation shaft 10 located withinthis swash-plate chamber 20, a swash-plate 21 is fixed with apredetermined inclination angle with respect to the rotation shaft 10such that this swash-plate rotates together with the rotation shaft 10.A plurality of locations in the circumferential direction of theswash-plate 21 and each of the pistons 17 are individually linked bymeans of a pair each of sliding shoes 22. Therefore, internal surfaces(mutually facing surfaces) of these individual sliding shoes 22 are madesmooth faces, and are slidingly contacted with a part near the outerdiameter on both side faces of the swash-plate 21 which are similarlysmooth faces. On the other hand, on the base end portion of therespective portions 17 (the end portion farther from the partition plate8; the left end portion in FIG. 6), is formed integral with each of thepistons 17, a connection portion 23 which together with the slidingshoes 22 and the swash-plate 21 constitutes a driving force transfermechanism. Moreover, a holding portion 24 for holding the pair ofsliding shoes 22 is formed on the respective connecting portions 17.

The outside end surface of each of the connecting portions 23, by meansof a guide surface (not shown in the figure), is allowed freedisplacement only in the axial direction (left-right direction in FIG.6) of the piston 17. Therefore, each of the pistons 17 is also fittedwithin the cylindrical bore 16 in such a way as to allow displacementonly in the axial direction (rotation is not possible). As a result,each of the connecting portions 23 pushes and pulls each of the pistons17 in the axial direction in accordance with the oscillating reciprocaldisplacement of the swash-plate 21 due to the rotation of the rotationshaft 10, and reciprocates each of the sliding portions 18 within thecylindrical bore 16 in the axial direction.

On the other hand, in the partition plate 8, which is sandwiched at thecontact portion between the main body 3 and the head case 4, forpartitioning the low pressure chamber 6, the high pressure chamber 7 andeach of the cylindrical bores 16, is formed penetrating in the axialdirection, an inlet 25 for communicating between the low pressurechamber and each cylindrical bore 16, and an outlet for communicatingbetween the high pressure chamber 7 and each cylindrical bore 16. Also,in the part of each of the cylindrical bores 16 which faces one end ofeach of the inlets 25, is provided a reed valve type inlet valve 27,which allows only flow of refrigerant vapor from the low pressurechamber 6 to each of the cylindrical bores 16. Also, in the part of thehigh pressure chamber 7 which faces the opening on the other end (rightside in FIG. 6) of the outlet 26, is provided a reed valve type outletvalve 28, which allows only flow of refrigerant vapor from thecylindrical bore 16 to the high pressure chamber 7. In this outlet valve28, a stopper 29, which restricts displacement in the direction awayfrom each of the outlet valve 26, is attached.

The rotation shaft 10 of the compressor 1 constructed in the abovemanner is driven by the propulsion engine of an automobile. Therefore,in the case of the example shown in the figure, on the periphery of asupport member, in other words a support cylinder 30, provided at thecenter of the outside surface (left side surface in FIG. 6) of theswash-plate case 5 constituting the casing 2, is rotationally supporteda driven pulley 31, by means of a double-row bearing. This driven pulley31 is constructed in an overall annular form with a C-shaped crosssection, and a solenoid 33, which is fixed to the outside surface of theswash-plate case 5, is provided within an internal cavity of the drivenpulley 31.

On the other hand, at an end portion of the rotation shaft, whichprotrudes from the support cylinder 30, is fixed a mounting bracket 34,and around the circumferential surface of this mounting bracket 34, issupported an annular plate of magnetic material, via a plate spring 36.This annular plate 35, when there is no current through the solenoid 33,is separated from the driven pulley 31 due to the elasticity of theplate spring 36, as shown in FIG. 6. However, when there is a currentthrough the solenoid 33, it is attracted towards this driven pulley 31,and hence allows the transmission of torque from this driven pulley 31to the rotation shaft 10. That is to say, the solenoid 33, the annularplate 35 and the plate spring 36, constitute an electromagnetic clutch37 for connecting and disconnecting the driven pulley 31 and therotation shaft 10. Also, between the driving pulley fixed to the end ofthe crank shaft of the propulsion engine and the driven pulley 31, isspanned an endless belt 38. Furthermore, in a state where the drivenpulley 31 and the rotation shaft 10 are connected by the electromagneticclutch 37, the rotation shaft 10 is rotated based on the rotation of theendless belt 38.

The operation of the swash-plate type compressor 1 formed in the abovemanner is as follows. That is to say, in order to perform cooling anddehumidification of the automobile interior, in the case of operating avapor compression type refrigerator, the rotation shaft 10 is rotated bythe propulsion engine, being the driving source. As a result, theswash-plate 21 rotates, and the sliding portions 18 constituting themultiple pistons 17 reciprocate within the respective cylindrical bores16. Furthermore, in accordance with such reciprocation of the slidingportions 18, the refrigerant vapor sucked in from the inlet port 9 issucked from the low pressure chamber 6 through each inlet 25 into thecompression chambers 19. This refrigerant vapor, after being compressedinside each of the compression chambers 19, is sent out to the highpressure chamber 7 via the outlets 26, and discharged from the outletport.

The compressor shown in FIG. 6 is one in which the inclination angle ofthe swash-plate with respect to the rotation shaft is unchangeable, andhence the refrigerant discharge volume is fixed. On the other hand, avariable displacement swash-plate type compressor in which theinclination angle of the swash-plate with respect to the rotation shaftcan be changed in order to change the discharge volume in accordancewith cooling load and the like, is conventionally widely known from, forexample, the disclosure of Japanese Unexamined Patent Publication No. H8-326655 and so on, and is commonly implemented. Moreover, as acompressor for a vapor compression type refrigerator constituting anautomobile air conditioning apparatus, the use of a scroll typecompressor is also being researched in some places. Furthermore, inrelation to a conventional compressor in which a piston is reciprocatedby means of a ball joint, this is still also being used in some places.

Whichever the structure of the compressor used, the compressorconstituting the automobile air conditioning apparatus is driven by theendless belt spanning between the driving pulley fixed to the end of thecrank shaft of the propulsion engine and the driven pulley provided onthe compressor side. Therefore, a radial load based on the tension forceof the endless belt, is exerted on the bearing which rotatably supportsthe driven pulley. In order to perform reliable power transmissionwithout slippage, between the endless belt and each of the pulleys, thetension force on the endless belt, in other words, the radial load,becomes correspondingly large. Therefore, as a bearing for supportingthe driven pulley, in order to support this large radial load, it isnecessary to use one with sufficient load capacity.

When the double row ball bearing 32 incorporated in the conventionalstructure shown in FIG. 6 is viewed from this perspective, the spacing Dof balls 39 arranged in a double row is large, and hence the structureis said to be one which can ensure sufficient load capacity. However,with the double row ball bearing 32, the dimensions in the axialdirection becomes bulky. On the other hand, recently, in considerationof the global environment, in an attempt to improve fuel efficiency ofautomobiles, miniaturization and lightening of automobile auxiliaryequipment such as the compressor is demanded. Furthermore, a demand hasalso arisen for shortening of the axial dimensions of rolling bearingsfor supporting driven pulleys incorporated into automobile auxiliaryequipment.

In response to such demands, as a rolling bearing for supporting thedriven pulley, the use of single row deep groove ball bearings and threepoint or four point contact type ball bearings is being researched.However, with such ball bearings, rigidity with respect to the load,mainly the moment load, exerted on the driven pulley, cannot be easilyensured, and it is difficult to ensure a sufficient low-vibrationproperty (propensity for not vibrating) or durability. That is to say,there are occasions where, though slight in magnitude, the moment loadfrom the driven pulley acts on the rolling bearing. However, rigidity ofthe single row deep groove type ball bearing with respect to the momentload is low. Also, regarding the three point to four point contact typeball bearing, though rigidity with respect to the moment load is higherthan the ordinary single row deep groove type ball bearing, there areoccasions where the rigidity is not always sufficient due to therelationships such as the magnitude of the tension force on the endlessbelt or the arrangement (eccentricity between the direction of radialload and the location of the ball bearing center). As a result,vibration as well as noise during the operation becomes more likely, andit is difficult to ensure durability.

The pulley support double row ball bearing of the present invention wasinvented in consideration of such circumstances.

RELATED ART

With such circumstances in mind, the present inventor first thought ofensuring the required rigidity by reducing the diameter of the balls andreducing the spacing between the balls arranged in double rows, as wellas supporting the driven pulley using a double row ball bearing withreduced dimensions related to the axial direction (Japanese PatentApplication No. 2002-24863, Japanese Patent Application No. 2002-97966).In the case of a pulley supporting double row ball bearing according tothese related inventions, one having an outer ring with an outerdiameter of less than 65 mm and a double row of outer ring raceways onthe inner circumferential surface is used. Also, an inner ring having adouble row of inner ring raceways on the outer circumferential surfaceis used. Moreover, balls with a diameter (major diameter) of less than 4mm are used, and several of these are provided so as to roll freelybetween each of the outer ring raceways and each of the inner ringraceways. Also, by using a retainer, each of the balls are held so as toallow free rolling. Moreover, a pair of seal rings is used to seal offthe openings on both sides of the inner space accommodating each of theballs between the inner circumferential surface of the outer ring andthe outer circumferential surface of the inner ring. Furthermore, thespacing between the balls, and the spacing between the balls and theseal ring are reduced, thus providing a double row ball bearing with anoverall width in the axial direction (approximately coinciding with theouter ring width and inner ring width) of less than 45% of the innerdiameter of this inner ring.

Also, in order to reduce the spacing between the balls, a crown shapedretainer made of synthetic resin is used for each of the retainers, andrims of each of the retainers are provided to oppose each other fromopposite sides (i.e. outsides in the axial direction, and sides opposedto the seal ring). Also, the distance between the rim of each of theretainers and the inside surface of the seal ring is reduced. However,again in this case, the distance between the rim of each of theretainers and the inside surface of each seal ring is ensured to be over13% of the diameter of each of the balls such that the filling amount ofthe grease within the inner space accommodating each of the balls,between both of the seal rings can be ensured.

According to the pulley support double row ball bearing associated withthe related invention, moment rigidity is ensured, while the widthrelated to the axial direction is reduced, and it is possible tocontribute to the realization of small and light automobile auxiliaryequipment, which produces low noise during operation.

DISCLOSURE OF THE INVENTION

Any of the pulley support double row ball bearings according to thepresent invention, in a similar manner to the aforementioned pulleysupport double row ball bearing associated with the related invention,is provided with: an outer ring with an outer diameter of less than 65mm and having a double row outer ring raceway on an innercircumferential surface; an inner ring having a double row inner ringraceway on an outer circumferential surface; balls with a diameter ofless than 4 mm, provided as several balls so as to be free rollingbetween the outer ring raceways and the inner raceways; a retainer whichholds these balls so as to be free rolling; and a seal ring, whichexists between the inner circumferential surface of the outer ring andthe outer circumferential surface of the inner ring, and seals offopenings on both ends of an inner space accommodating the balls.Furthermore, a width of the bearing related to the axial direction isless than 45% of the inner diameter of the inner ring, and by externallyfitting this inner ring to a support member and internally fitting theouter ring to a pulley, the pulley is rotatably supported on theperiphery of this support member.

Particularly, in a first aspect of the pulley support double row ballbearing of this invention, a portion near an inner circumference of therespective seal rings and both end surfaces in the axial direction ofthe inner ring overlap when viewed from the axial direction, so that awidth in the radial direction of an overlap section is 25% or more thana diameter of the respective balls. Also, a plurality of protrusions areformed all around a circumference on an inside surface at a portion nearan inner circumference of the respective seal rings, and a tip edge ofat least one of the protrusions comes in sliding contact with the endsurfaces in the axial direction of the inner ring.

Moreover, in a second aspect of the pulley support double row ballbearing of this invention, a portion near an inner circumference of therespective seal rings and both end surfaces in the axial direction ofthe inner ring overlap when viewed from the axial direction, so that awidth in the radial direction of an overlap section is 25% or more thana diameter of the respective balls. Also, one or more protrusions areformed all around a circumference on a side surface at a portion near aninner circumference of the respective seal rings, and a tip edge of atleast one of the protrusions comes in sliding contact all the way aroundthe circumference with a part of the surface of the inner ring. Togetherwith this, another portion near the inner circumference of therespective seal rings, not being a protusion in sliding contact, comesclose to and faces the other part of the surface of the inner ring, sothat labyrinth seals are formed.

In a third aspect of the double row ball bearing for pulley support ofthis invention, the seal rings comprise an elastic material having aShore hardness of 60 to 80 and reinforced by a metal core. The width inthe radial direction of a deformed section of the elastic material thatprotrudes inward in the radial direction from the inner edge of themetal core is 40% or more than the diameter of the respective balls, andthe thickness of the thinnest area of this deformed section, which islocated in the middle in the radial direction, is 0.4 mm or more.

In a fourth aspect of the double row ball bearing for pulley support ofthis invention, the seal rings comprise an elastic material that isreinforced by a metal core. Also, an inner diameter of this metal coreis less than an outer diameter of the inner ring.

Furthermore, in a fifth aspect of the double row ball bearing for pulleysupport of this invention, the seal rings comprise an elastic materialthat is reinforced by a metal core. Also, a position in the axialdirection of the center of gravity of the deformed section of theelastic material that protrudes inward in the radial direction from theinner edge of the metal core is located more adjacent to the side of thesliding contact between the tip edge of the seal ring and part of thesurface of the inner ring than the position of the center of deformationof this deformed section.

In the embodiments of this invention, the aforementioned aspects can beembodied singly or in appropriate combinations. It is also possible tocombine all of the aspects described above.

In the case of the double row ball bearing for pulley support of thisinvention constructed as described above, the seal rings make a goodseal at the openings on both ends, and it is possible to maintainexcellent durability even under severe conditions.

First, in the case of the double row ball bearing for pulley support ofthe first and second aspects, it is possible to maintain the width inthe axial direction of the overlap sections of the inner edge sectionsof the seal rings and the end surfaces in the axial direction of theinner ring, and since there is a plurality of protrusions in thisoverlap section (in the case of the first aspect), or protrusions andlabyrinth seals (in the case of the second aspect), it is possible tohave a good seal in the overlap sections.

Moreover, in the case of the double row ball bearing for pulley supportof the third and fourth aspects, it is possible to maintain the rigidityof the elastic material of the seal rings, as well as maintain thesurface pressure at the areas of sliding contact between the tip edgesof this elastic material and parts of the surface of the inner ring,therefore it is possible for the seal rings to make a good seal.

Furthermore, in the case of the double row ball bearing for pulleysupport of the fifth aspect, the centrifugal force that acts on the seallips during operation acts in a direction that pushes the seal lipsformed on the tip edges of the seal rings toward parts of the surface ofthe inner ring. As a result, it is possible to maintain the surfacepressure at the areas of sliding contact between the tip edges of thiselastic material and parts of the surface of the inner ring, thereforeit is possible for the seal rings to make a good seal.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a cross-sectional view that shows a first example of anembodiment of the invention.

FIG. 2 is an enlarged view of the upper right area in FIG. 1.

FIG. 3 is a drawing similar to FIG. 2 that shows a second example of anembodiment of the invention.

FIG. 4 is a partial cross-sectional drawing that corresponds to thelower right area of FIG. 2 and shows a third example of an embodiment ofthe invention.

FIG. 5 is a partial cross-sectional view that corresponds to the lowerright area in FIG. 2 and shows a fourth example of an embodiment of theinvention.

FIG. 6 is a cross-sectional view showing an example of a prior knowncompressor.

FIG. 7 is a drawing showing the view in the direction of the arrow A inFIG. 6.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

FIGS. 1 to 2 show a first embodiment of the invention and correspond toa first, second, third and fifth aspect of the invention. In FIG. 1 andFIG. 2, (and FIG. 3 to be described later) the proportions of thedimensions of each of the parts are drawn based on the actualproportions. In the case of the pulley support double row ball bearing32 a of this embodiment, an outer race with an outer diameter D₄₀ (seeFIG. 1) of 65 mm or less (D₄₀≦65 mm) and having double rows of outerraceways 41 formed around its inner circumferential surface is used asthe outer ring 40. Also an inner ring having double rows of innerraceways 43 formed around its outer circumferential surface is used asthe inner ring 42. Moreover, balls 44 having a diameter (outer diameter)D₄₄ (see FIG. 1 and 2) of 4 mm or less (D₄₄≦4 mm) are used (for allpractical purposes, the balls used are 3 to 4 mm), and a plurality ofballs is located between each of the outer raceways 41 and innerraceways 43 such that they can roll freely. Also, a pair of retainers 45holds the balls 44 such that they can roll freely, and a pair of sealrings 46 covers the opening on both ends of the internal space betweenthe inner circumferential surface of the outer ring 40 and the outercircumferential surface of the inner ring 42 where the balls 44 arelocated. The width W₃₂ of the double row ball bearing 32 a (FIG. 1) is45% or less than the inner diameter R₄₂ (FIG. 1) of the inner ring 42(W₃₂≦0.45 R₄₂).

Throughout all of the drawings the same reference numbers will be usedfor the same components.

Each of the seal rings 46 comprises a circular-shaped metal core 48 thatis made of metal such as steel sheet and is reinforced by an elasticmaterial 49 such as nitrile rubber, heat-resistant nitrile rubber,acrylic rubber or fluororubber, and it has a complete circular shape.Also, an elastic material having a Shore hardness (HS) within the range60 to 80 is used as this elastic material 49. Moreover, the outer edgeof this elastic material 49 has an attachment section 50 that protrudesoutward in the radial direction further than the outer edge of the metalcore 48, and this attachment section 50 is attached to an attachmentgroove 51 that is formed around the inner circumferential surface onboth ends of the outer ring 40. Also, the inner half in the radialdirection of this elastic material 49 protrudes inward in the radialdirection further than the metal core 48 and functions as a seal lip 52.

In the case of this embodiment, when viewed from the axial direction,this seal lip 52 overlaps the inner ring 42 by an amount that is 25% ormore than the diameter D₄₄ of the balls 44. In other words, when theinner diameter of the seal lip 52 is taken to be R₅₂ (FIG. 1) and theouter diameter of the inner ring 42 is taken to be D₄₂ (FIG. 1), theinside surface of the seal lip 52 and the surface on both ends in theaxial direction of the inner ring 42 face each other in the ring-shapedarea such that the width in the radial direction is ½ the differencebetween the outer diameter D₄₂ and the inner diameter R₅₂ {(D₄₂−R₅₂)/2}.In this embodiment, the width of this ring-shaped area 25% or greaterthan the diameter D₄₄ of the balls 44 {(D₄₂−R₅₂)/2 ≧0.25 D₄₄}, and morepreferably greater than 35% the diameter D₄₄. The maximum value for thewidth of this ring-shaped area is not particularly set, however, fromthe aspect of preventing the bearing from becoming large, it is notpractical for it to be greater than the diameter D₄₄. In order to have acompact pulley support double row ball bearing, it is best to keep thewidth less than 80% and more preferably less than 70% of the diameterD₄₄.

On the other hand, a first to third protrusion 53 to 55 are formed inorder from the inner side in the radial direction all the way around thecircumference on the inside surface of the seal lip 52 such that theyare concentric with each other. With the outer edge of each seal ring 46attached to the attachment groove 51, of these protrusions 53 to 55, thetip edge of the first protrusions 53 that are located the furthestinward in the radial direction comes in sliding contact along the entirecircumference with the surfaces 56 on both ends in the axial directionof the inner ring 42. On the other hand, the tip edges of the secondprotrusions 54 that are located in the middle and the third protrusions55 that are located the furthest outward in the radial direction comevery close to and face both end surfaces 56 or the corner sections 57that are located in the connecting areas between the outercircumferential surface of the inner ring 42 and both end surfaces 56,to form labyrinth seals in those areas.

Moreover, the width W₅₂ in the radial direction of the seal lips (FIG.2) as a deformed section in the third aspect of the invention, is takento be 40% or more of the diameter D₄₄ of the balls (W₅₂≧0.4 D₄₄). Also,the thickness T₅₂ (FIG. 2) at the thinnest section that is located inthe middle in the radial direction of the seal lip 52 (it is notnecessary to be in the middle) is taken to be 0.4 mm or more (T₅₂>0.4mm). The upper limit for the thickness T₅₂ of this section in notparticularly limited, however, the rigidity of the seal lip 52 shouldnot be greater than necessary, and in order to keep the material costsdown, it is preferred that it be 0.6 mm or less and even more preferablyless than 0.5 mm. Moreover, it is preferred that the width W₅₂ of theseal lip be 60% or less than that of the diameter D₄₄, and morepreferably less than 50%.

Furthermore, in this embodiment, in regards to the position in the axialdirection (left-right direction in FIG. 1 and FIG. 2), the deformedsections of the elastic material 49 that protrude inward in the radialdirection from the inner edges of the metal cores 48, or in other words,the position of the center of gravity G (FIG. 2) of the seal lips islocated further on the side where there is sliding contact between thetip edges of the seal lips 52 (first and second protrusions 53, 54) andboth end surfaces in the axial direction of the inner ring 42 than thecenter of deformation of the deformed sections of the seal lips 52. Inother words, the shape and dimensions of each of the parts is regulatedsuch that the seal lip 52 elastically deforms around the inner edge ofthe metal core 48 or the thinnest section of the seal lip 52, and whenthe outer edge section of each seal ring 46 is attached to theattachment groove 51, the position of the center of gravity G is locatedfurther on the side of both end surfaces 56 of the inner ring than thecenter of the elastic deformation.

In the case of the pulley support double row ball bearing 32 a of thisembodiment constructed as described above, there is good sealperformance by both seal rings 46 that are located at the openings onboth ends, so even when operating under severe conditions, it ispossible to maintain excellent durability. In other words, in the caseof the pulley support double row ball bearing 32 a of this embodiment,the width in the radial direction of the overlap portions in the axialdirection of the inner edges of the seal rings 46 and both end surfacesin the axial direction of the inner ring 42 is maintained. Also, the tipedges of the first and second protrusions 53, 54 located in the overlapsections come in sliding contact all along their circumference with bothend sections 56, and labyrinth seals 58 are formed between the thirdprotrusions 55 and the corner sections 57. Therefore, there is a goodseal in the overlap sections.

The elastic material 49 of the seal rings 46 is made of rubber having aShore hardness of 60 to 80, and the thickness T₅₂ of the middle section(the thinnest section) of the seal lip 52 that is formed on the innerdiameter half of this elastic material 49 is maintained at 0.4 mm orgreater, so it is possible to maintain the rigidity of the seal lip 52.As a result, it is possible to maintain surface pressure at the areas ofsliding contact between the tip edges of the first and sectionprotrusions 53, 54 that are formed on the inside surface of the seallips 52 and both end surfaces 56 in the axial direction of the innerring 42, and thus it is possible to make a good seal by the seal rings46.

Furthermore, in the case of the pulley support double row ball bearing32 a of this embodiment, by properly regulating the position of thecenter of gravity G, the centrifugal force that is applied to the seallips 52 of each of the seal rings 46 during operation acts in adirection such that the seal lips 52 are elastically deformed towardboth end surfaces 56 of the inner ring 42. As a result, it is possibleto maintain surface pressure at the areas of sliding contact between thetip edges of the first and section protrusions 53, 54 and both endsurfaces 56 in the axial direction of the inner ring 42 (it is possibleto prevent a drop in surface pressure during high-speed rotation), andthus it is possible to make a good seal by the seal rings 46.

It is not directly related to the invention, however, each of the outerraceways and inner raceways are deep-groove type tracks, and from theaspect of maintaining the moment rigidity of the pulley support doublerow ball bearing, it is preferred that the groove depth D₄₁ (FIG. 2) ofeach of the outer raceways 41 be kept at 18% or more than the diameterD₄₄ of the balls 44, and that the groove depth D₄₃ of the inner raceway43 (FIG. 2) be kept at 20% or more than the diameter D₄₄ (D₄₁≧0.18 D₄₄,D₄₃≧0.20 D₄₄). However, the total depth of both grooves D₄₁, D₄₃ is keptat 42% or less than the diameter D₄₄ {(D₄₁+D₄₃)≦0.42 D₄₄}, so it ispossible to install the balls 44 between both raceways 41, 43.

Next, FIG. 3 shows a second example of an embodiment of the inventionthat corresponds to a second, third, forth and fifth aspect of theinvention. In the case of the double row ball bearing 32 b of thisembodiment, the inner diameter R₄₈ of the metal core 48 of the seal ring46 a is taken to be less than the outer diameter D₄₂ of the inner ring42 (R₄₈<D₄₂). Also, only a first protrusion 53 is formed on the inneredge of the inside surface of the seal lip 52 a that is formed on theinner half in the radial direction of the elastic material 49 a and athird protrusion 55 is formed on the base end (the outer end in theradial direction), and the second protrusion 54 in the middle is omitted(see FIG. 1 and FIG. 2). Instead, a fourth protrusion 59 is formedaround the inner edge on the outside surface of the seal lip 52 a.

When the outer edges of the seal rings 46 a are attached to theattachment grooves 51 that are formed around the inner circumferentialsurfaces on both ends of the outer ring 40, only the tip edges of thefirst protrusions 53 come in sliding contact all the way around theircircumference with the end surfaces 56 in the axial direction of theinner ring 42. At the same time, the tip edges of the third protrusions55 come close to and face the corner sections 57, and the fourthprotrusions 59 come close to and face protrusions 60 that are formedaround the outer circumferential surfaces on the ends of the inner ring42 to form labyrinth seals 58, 61 in those areas.

In the case of the embodiment described above, by keeping the innerdiameter R₄₈ of the metal core 48 a less then the outer diameter D₄₂ ofthe inner ring 42, the rigidity of the seal lips 52 a is increased, andit becomes easy to maintain surface pressure at the areas of slidingcontact between the tip edges of the first protrusions 53 and the endsurfaces 56 in the axial direction of the inner ring 42. Furthermore,since the areas of sliding contact are located between pairs oflabyrinth seals 58, 61, it is possible to sufficiently maintain a goodseal at the openings on both ends of the pulley support double row ballbearing 32 b.

Next, FIG. 4 shows a third embodiment of the invention that correspondsto the second and third aspects of the invention. In this embodiment,the thickness of the section of the inner half of the elastic material49 b of the seal ring 46 b that protrudes inward in the radial directionfurther than the inner edge of the metal core 48 suddenly decreases inthe middle section as it goes from the outside inward in the radialdirection. Also, an inside stepped section 62 and an outside steppedsection 63 are formed respectively on the inside surface and outsidesurface of the middle section in the radial direction. When the outeredge of the seal ring 46 b is attached around the inner circumferentialsurface on the end of the outer ring 40 (see FIGS. 1 to 3), the insidestepped section 62 comes close to and faces the corner section 57 aroundthe outer circumferential surface of the inner ring 42, and the outsidestepped section comes close to and faces the protrusion 60 torespectively form labyrinth seals 58, 61 in those areas. The tip edge ofthe protrusion 64 that is formed around the inner edge of the outsidesurface of the seal lip 52 a and that is located further inward in theradial direction than both of the stepped sections 62, 63 comes insliding contact all the way around its circumference with the insidesurface of the protrusion 60 that is part of the surface of the innerring 42. In the case of this embodiment as well, similar to the secondembodiment described above, the area of sliding contact is locatedbetween a pair of labyrinth seals 58, 61, so it is possible tosufficiently maintain a good seal at the openings on both ends of thepulley support double row ball bearing.

Next, FIG. 5 shows a fourth embodiment of the invention that correspondsto the second and third aspects of the invention. In the case of thisembodiment, the tip edge of the protrusion 65 that is formed all the wayaround the circumference on the inside surface of the middle section ofthe seal lip 52 a comes close to and faces the end surface 56 in theaxial direction of the inner ring 42 and forms a labyrinth seal in thatarea. In this embodiment, compared with the third embodiment describedabove, the number of labyrinth seals is increased by this labyrinth seal66, so it is possible to further improve the seal performance at theopenings on both ends of the pulley support double row ball bearing.

INDUSTRIAL APPLICABILITY

The pulley support double row ball bearing of this invention isconstructed and functions as described above, so even when used undersevere conditions, it is possible to effectively prevent foreign mattersuch as muddy water from getting inside the bearing, and it cancontribute to making various auxiliary equipment of an automobile, suchas a compressor, more compact and lightweight, while at the same timemaintaining durability.

1. A pulley support double row ball bearing comprising: an outer ringhaving an outer diameter of 65 mm or less and a double row racewayformed on an inner circumferential surface thereof; an inner ring havinga double rowraceway formed on an outer circumferential surface thereof;a plurality of balls each 4 mm or less in diameter, and each retained bya retainer between the outer and inner raceways such that they rollfreely; and seal rings that seal openings on both ends of an internalspace between the inner circumferential surface of the outer ring andthe outer circumferential surface of the inner ring where the pluralityof balls are disposed; wherein an axial width of the bearing does notexceed 45% of an inner diameter of the inner ring, and by fitting theinner ring around a support member and fitting the outer ring inside apulley, the pulley is supported such that it rotates freely around thesupport member; and wherein a portion of each seal ring near an innercircumference thereof and a corresponding axial end surface of the innerring overlap when viewed from the axial direction, so that a width in aradial direction of an overlap section is between 25% and 80% of adiameter of one of the plurality of balls; and wherein each seal ringincludes a plurality of protrusions formed circumferentially on aninside surface at a portion near an inner circumference of the seal ringsuch that a tip edge of at least one of the plurality of protrusionscomes in sliding contact with the corresponding axial end surface of theinner ring.
 2. A pulley support double row ball bearing described inclaim 1, wherein each seal ring comprises an elastic material reinforcedby a metal core, and a position in an axial direction of a center ofgravity of a deformed section of the elastic material that protrudesinward in a radial direction from an inner edge of the metal core islocated more adjacent to a side where the tip edge of the seal ring andpart of the surface of the inner ring come into sliding contact, thanthe position of the center of deformation of the deformed section.
 3. Apulley support double row ball bearing comprising: an outer ring havingan outer diameter of 65 mm or less and a double row raceway formed on aninner circumferential surface thereof; an inner ring having a doublerowraceway formed on an outer circumferential surface thereof; aplurality of balls each 4 mm or less in diameter, and each retained by aretainer between the outer and inner raceways such that they rollfreely; and seal rings that seal openings on both ends of an internalspace between the inner circumferential surface of the outer ring andthe outer circumferential surface of the inner ring where the pluralityof balls are disposed; wherein an axial width of the bearing does notexceed 45% of an inner diameter of the inner ring, and by fitting theinner ring around a support member and fitting the outer ring inside apulley, the pulley is supported such that it rotates freely around thesupport member; and wherein a portion of each seal ring near an innercircumference thereof and a corresponding axial end surface of the innerring overlap when viewed from the axial direction, so that a width in aradial direction of an overlap section is between 25% and 80% of adiameter of one of the plurality of balls; and wherein each seal ringincludes one or more protrusions formed circumferentially on an insidesurface at a portion near an inner circumference such that a tip edge ofat least one of the protrusions comes in sliding contact with a part ofthe surface of the inner ring all the way around the circumference; andwherein each seal ring includes portions not in sliding contact with theinner ring near the inner circumference of the respective seal rings,said portions disposed close to and facing a portion of the surface ofthe inner ring not in sliding contact with the protrusions, so thatlabyrinth seals are formed.
 4. A pulley support double row ball bearingcomprising: an outer ring having an outer diameter of 65 or less and adouble row raceway formed on an inner circumferential surface thereof;an inner ring having a double rowraceway formed on an outercircumferential surface thereof; a plurality of balls each 4 mm or lessin diameter, and each retained by a retainer between the outer and innerraceways such that they roll freely; and seal rings that seal openingson both ends of an internal space between the inner circumferentialsurface of the outer ring and the outer circumferential surface of theinner ring where the plurality of balls are disposed; wherein an axialwidth of the bearing does not exceed 45% of an inner diameter of theinner ring, and by fitting the inner ring around a support member andfitting the outer ring inside a pulley, the pulley is supported suchthat it rotates freely around the support member; and wherein each sealring comprises an elastic material having a Shore hardness of 60 to 80and reinforced by a metal core, and the width in a radial direction of adeformed section of the elastic material that protrudes inward in theradial direction from an inner edge of the metal core is 40% or morethan the diameter of one of the plurality of balls, and a thickness of athinnest area of the deformed section, which is located in a middle inthe radial direction of the deformed section, is between 0.4 mm and 0.6mm.